Internal combustion engines are any of a group of devices in which the reactants of combustion, e.g., oxidizer and fuel, and the products of combustion serve as the working fluids of the engine. The basic components of an internal combustion engine are well known in the art and include the engine block, cylinder head, cylinders, pistons, valves, crankshaft and camshaft. The cylinder heads, cylinders and tops of the pistons typically form combustion chambers into which fuel and oxidizer (e.g., air) is introduced and combustion takes place. Such an engine gains its energy from the heat released during the combustion of the non-reacted working fluids, e.g., the oxidizer-fuel mixture. This process occurs within the engine and is part of the thermodynamic cycle of the device. In all internal combustion engines, useful work is generated from the hot, gaseous products of combustion acting directly on moving surfaces of the engine, such as the top or crown of a piston. Generally, reciprocating motion of the pistons is transferred to rotary motion of a crankshaft via connecting rods.
Internal combustion (IC) engines can be categorized into spark ignition (SI) and compression ignition (CI) categories. SI engines, i.e. typical gasoline engines, use a spark to ignite the air-fuel mixture, while the heat of compression ignites the air fuel mixture in CI engines, i.e., typically diesel engines.
The most common internal-combustion engine is the four-stroke cycle engine, a conception whose basic design has not changed for more than 100 years old. This is because of its outstanding performance as a prime mover in the ground transportation industry. In a four-stroke cycle engine, power is recovered from the combustion process in four separate piston movements (strokes) of a single piston. For purposes herein, a stroke is defined as a complete movement of a piston from a top dead center position to a bottom dead center position or vice versa. Accordingly, a four-stroke cycle engine is defined herein to be an engine which requires four complete strokes of one or more pistons for every power stroke, i.e. for every stroke that delivers power to a crankshaft.
Referring to FIGS. 1–4, an exemplary embodiment of a prior art four stroke cycle internal combustion engine is shown at 10. For purposes of comparison, the following four FIGS. 1–4 describe what will be termed a prior art “standard engine” 10. As will be explained in greater detail hereinafter, this standard engine 10 is an SI engine with a 4 inch diameter piston, a 4 inch stroke and an 8 to 1 compression ratio. The compression ratio is defined herein as the maximum volume of a predetermined mass of an air-fuel mixture before a compression stroke, divided by the volume of the mass of the air-fuel mixture at the point of ignition. For the standard engine, the compression ratio is substantially the ratio of the volume in cylinder 14 when piston 16 is at bottom dead center to the volume in the cylinder 14 when the piston 16 is at top dead center.
The engine 10 includes an engine block 12 having the cylinder 14 extending therethrough. The cylinder 14 is sized to receive the reciprocating piston 16 therein. Attached to the top of the cylinder 14 is the cylinder head 18, which includes an inlet valve 20 and an outlet valve 22. The cylinder head 18 cylinder 14 and top (or crown 24) of the piston 16 form a combustion chamber 26. On the inlet stroke (FIG. 1), a fuel air mixture is introduced into the combustion chamber 26 through an intake passage 28 and the inlet valve 20, wherein the mixture is ignited via spark plug 30. The products of combustion are later exhausted through outlet valve 22 and outlet passage 32 on the exhaust stroke (FIG. 4). A connecting rod 34 is pivotally attached at its top distal end 36 to the piston 16. A crankshaft 38 includes a mechanical offset portion called the crankshaft throw 40, which is pivotally attached to the bottom distal end 42 of connecting rod 34. The mechanical linkage of the connecting rod 34 to the piston 16 and crankshaft throw 40 serves to convert the reciprocating motion (as indicated by arrow 44) of the piston 16 to the rotary motion (as indicated by arrow 46) of the crankshaft 38. The crankshaft 38, is mechanically linked (not shown) to an inlet camshaft 48 and an outlet camshaft 50, which precisely control the opening and closing of the inlet valve 20 and outlet valve 22 respectively.
The cylinder 14 has a centerline (piston-cylinder axis) 52, which is also the centerline of reciprocation of the piston 16. The crankshaft 38 has a center of rotation (crankshaft axis) 54. For purposes of this specification, the direction of rotation 46 of the crankshaft 38 will be in the clockwise direction as viewed by the reader into the plane of the paper. The centerline 52 of the cylinder 14 passes directly through the center of rotation 54 of the crankshaft 38.
Referring to FIG. 1, with the inlet valve 20 open, the piston 16 first descends (as indicated by the direction of arrow 44) on the intake stroke. A predetermined mass of an explosive mixture of fuel (gasoline vapor) and air is drawn into the combustion chamber 26 by the partial vacuum thus created. The piston continues to descend until it reaches its bottom dead center (BDC), the point at which the piston is farthest from the cylinder head 18.
Referring to FIG. 2, with both the inlet 20 and outlet 22 valves closed, the mixture is compressed as the piston 16 ascends (as indicated by the direction of arrow 44) on the compression stroke. As the end of the stroke approaches top dead center (TDC), i.e., the point at which the piston 16 is closest to the cylinder head 18, the volume of the mixture is compressed to one eighth of its initial volume (due to an 8 to 1 compression ratio). The mixture is then ignited by an electric spark from spark plug 30.
Referring to FIG. 3, the power stroke follows with both valves 20 and 22 still closed. The piston 16 is driven downward (as indicated by arrow 44) toward bottom dead center (BDC), due to the expansion of the burned gas pressing on the crown 24 of the piston 16. Since the spark plug 30 is fired when the piston 16 is at or near TDC, i.e. at its firing position, the combustion pressure (indicated by arrow 56) exerted by the ignited gas on the piston 16 is at its maximum at this point. This pressure 56 is transmitted through the connecting rod 34 and results in a tangential force or torque (as indicated by arrow 58) on the crankshaft 38.
When the piston 16 is at ifs firing position, there is a significant clearance distance 60 between the top of the cylinder 14 and the crown 24 of the piston 16. Typically, the clearance distance is between 0.5 to 0.6 inches. For the standard engine 10 illustrated the clearance distance is substantially 0.571 inches. When the piston 16 is at its firing position conditions are optimal for ignition, i.e., optimal firing conditions. For purposes of comparison, the firing conditions of this engine 10 exemplary embodiment are: 1) a 4 inch diameter piston, 2) a clearance volume of 7.181 cubic inches, 3) a pressure before ignition of approximately 270 pounds per square inch absolute (psia), 4) a maximum combustion pressure after ignition of approximately 1200 psia and 5) operating at 1400 RPM.
This clearance distance 60 corresponds typically to the 8 to 1 compression ratio. Typically, SI engines operate optimally with a fixed compression ratio within a range of about 6.0 to 8.5, while the compression ratios of CI engines typically range from about 10 to 16: The piston's 16 firing position is generally at or near TDC, and represents the optimum volume and pressure for the fuel-air mixture to ignite. If the clearance distance 60 were made smaller, the pressure would increase rapidly.
Referring to FIG. 4, during the exhaust stroke the ascending piston 16 forces the spent products of combustion through the open outlet (or exhaust) valve 22. The cycle then repeats itself. For this prior art four stoke cycle engine 10, four stokes of each piston 16, i.e. inlet, compression, power and exhaust, and two revolutions of the crankshaft 38 are required to complete a cycle, i.e. to provide one power stroke.
Problematically, the overall thermodynamic efficiency of the standard four stroke engine 10 is only about one third (⅓). That is ⅓ of the work is delivered to the crankshaft, ⅓ is lost in waste heat, and ⅓ is lost out of the exhaust.
As illustrated in FIGS. 3 and 5, one of the primary reasons for this low 20 efficiency is the fact that peak torque and peak combustion pressure are inherently locked out of phase. FIG. 3 shows the position of the piston 16 at the beginning of a power stroke, when the piston 16 is in its firing position at or near TDC. When the spark plug 30 fires, the ignited fuel exerts maximum combustion pressure 56 on the piston 16, which is transmitted through the connecting rod 34 to the crankshaft throw 40 of crankshaft 38. However, in this position, the connecting rod 34 and the crankshaft throw 40 are both nearly aligned with the centerline 52 of the cylinder 14. Therefore, the torque 58 is almost perpendicular to the direction of force 56, and is at its minimum value. The crankshaft 38 must rely on momentum generated from an attached flywheel (not shown) to rotate it past this position.
Referring to FIG. 5, as the ignited gas expands in the combustion chamber 26, the piston 16 descends and the combustion pressure 56 decreases. However, as the crankshaft throw 40 rotates past the centerline 52 and TDC, the resulting tangential force or torque 58 begins to grow. The torque 58 reaches a maximum value when the crankshaft throw 40 rotates approximately 30 degrees past the centerline 52. Rotation beyond that point causes the pressure 56 to fall off so much that the torque 58 begins to decrease again, until both pressure 56 and torque 58 reach a minimum at BDC. Therefore, the point of maximum torque 58 and the point of maximum combustion pressure 56 are inherently locked out of phase by approximately 30 degrees.
Referring to FIG. 6, this concept can be further illustrated. Here, a graph of tangential force or torque versus degrees of rotation from TDC to BDC is shown at 62 for the standard prior art engine 10. Additionally, a graph of combustion pressure versus degrees of rotation from TDC to BDC is shown at 64 for engine 10. The calculations for the graphs 62 and 64 were based on the standard prior art engine 10 having a four inch stroke, a four inch diameter piston, and a maximum combustion pressure at ignition of about 1200 PSIA. As can be seen from the graphs, the point of maximum combustion pressure 66 occurs at approximately 0 degrees from TDC and the point of maximum torque 68 occurs approximately 30 degrees later when the pressure 64 has been reduced considerably. Both graphs 62 and 64 approach their minimum values at BDC, or substantially 180 degrees of rotation past TDC.
An alternative way of increasing the thermal dynamic efficiency of a four stoke cycle engine is to increase the compression ratio of the engine. However, automotive manufactures have found that SI engines typically operate optimally with a compression ratio within a range of about 6.0 to 8.5, while CI engines typically operate best within a compression ratio range of about 10 to 16. This is because as the compression ratios of SI or CI engines increase substantially beyond the above ranges, several other problems occur which outweigh the advantages gained. For example, the engine must be made heavier and bulkier in order to handle the greater pressures involved. Also problems of premature ignition begin to occur, especially in SI engines.
Many rather exotic early engine designs were patented. However, none were able to offer greater efficiencies or other significant advantages, which would replace the standard engine 10 exemplified above. Some of these early patents included: U.S. Pat. Nos. 848,029; 939,376; 1,111,841; 1,248,250; 1,301,141; 1,392,359; 1,856,048; 1,969,815; 2,091,410; 2,091,411; 2,091,412; 2,091,413; 2,269,948; 3,895,614; British Patent No. 299,602; British Patent No. 721,025 and Italian Patent No. 505,576. In particular the U.S. Pat. No. 1,111,841 to Koenig disclosed a prior art split piston/cylinder design in which an intake and compression stroke was accomplished in a compression piston 12/cylinder 11 combination, and a power and an exhaust stroke was accomplished in an engine piston 7/cylinder 8 combination. Each piston 7 and 12 reciprocates along a piston cylinder axis which intersected the single crankshaft 5 (see FIG. 3 therein). A thermal chamber 24 connects the heads of the compression and engine cylinders, with one end being open to the engine cylinder and the other end having a valued discharge port 19 communicating with the compressor cylinder. A water cooled heat exchanger 15 is disposed at the top of the compressor cylinder 11 to cool the air or air/fuel mixture as it is compressed. A set of spaced thermal plates 25 are disposed within the thermal chamber 24 to re-heat the previously cooled compressed gas as it passes through.
It was thought that the engine would gain efficiency by making it easier to compress the gas by cooling it. Thereafter, the gas was re-heated in the thermal chamber in order to increase its pressure to a point where efficient ignition could take place. Upon the exhaust stroke, hot exhaust gases were passed back through the thermal chamber and out of an exhaust port 26 in an effort to re-heat the thermal chamber.
Unfortunately, transfer of gas in all prior art engines of a split piston design always requires work, which reduces efficiency. Additionally, the added expansion from the thermal chamber to the engine cylinder of Koenig also reduced compression ratio. The standard engine 10 requires no such transfer process and associated additional work. Moreover, the cooling and re-heating of the gas, back and forth through the thermal chamber did not provide enough of an advantage to overcome the losses incurred during the gas transfer process. Therefore, the Koenig patent lost efficiency and compression ratio relative to the standard engine 10.
For purposes herein, a crankshaft axis is defined as being offset from the piston cylinder axis when the crankshaft axis and the piston-cylinder axis do not intersect. The distance between the extended crankshaft axis and the extended piston-cylinder axis taken along a line drawn perpendicular to the piston cylinder axis is defined as the offset. Typically, offset pistons are connected to the crankshaft by well-known connecting rods and crankshaft throws. However, one skilled in the art would recognize that offset pistons may be operatively connected to a crankshaft by several other mechanical linkages. For example, a first piston may be connected to a first crankshaft and a second piston may be connected to a second crankshaft, and the two crankshafts may be operatively connected together through a system of gears. Alternatively, pivoted lever arms or other mechanical linkages may be used in conjunction with, or in lieu of, the connecting rods and crankshaft throws to operatively connect the offset pistons to the crankshaft.
Certain technology relating to reciprocating piston internal combustion engines in which the crankshaft axis is offset from, i.e., does not intersect with, the piston-cylinder axes is described in U.S. Pat. Nos. 810,347; 2,957,455; 2,974,541; 4,628,876; 4,945,866; and 5,146,884; in Japan patent document 60-256,642; in Soviet Union patent document 1551-880-A; and in Japanese Society of Automotive Engineers (JSAE) Convention Proceedings, date 1996, issue 966, pages 129–132. According to descriptions contained in those publications, the various engine geometries are motivated by various considerations, including power and torque improvements and friction and vibration reductions. Additionally, in-line, or straight engines in which the crankshaft axis is offset from the piston axes were used in early twentieth century racing engines.
However, all of the improvements gained were due to increasing the torque angles on the power stroke only. Unfortunately, as will be discussed in greater detail hereinafter, the greater the advantage an offset was to the power stroke was also accompanied by an associated increasing disadvantage to the compression stroke. Therefore, the degree of offset quickly becomes self limiting, wherein the advantages to torque, power, friction and vibration to the power stroke do not out weigh the disadvantages to the same functions on the compression stroke. Additionally, no advantages were taught or discussed regarding offsets to optimize the compression stroke.
By way of example, a recent prior art attempt to increase efficiency in a standard engine 10 type design through the use of an offset is disclosed in U.S. Pat. No. 6,058,901 to Lee. Lee believes that improved efficiency will result by reducing the frictional forces of the piston rings on the side walls over the full duration of two revolutions of a four stroke cycle (see Lee, column 4, lines 1016). Lee attempts to accomplish this by providing an offset cylinder, wherein the timing of combustion within each cylinder is controlled to cause maximum combustion pressure to occur when an imaginary plane that contains both a respective connection axis of a respective connecting rod to the respective piston and a respective connection axis of the connecting rod to a respective throw of the crankshaft is substantially coincident with the respective cylinder axis along which the piston reciprocates.
However, though the offset is an advantage during the power stroke, it becomes a disadvantage during the compression stroke. That is, when the piston travels from bottom dead center to top dead center during the compression stroke, the offset piston-cylinder axis creates an angle between the crankshaft throw and connecting rod that reduces the torque applied to the piston. Additionally, the side forces resulting from the poor torque angles on the compression stroke actually increase wear on the piston rings. Accordingly, a greater amount of power must be consumed in order to compress the gas to complete the compression stroke as the offset increases. Therefore, the amount of offset is severely limited by its own disadvantages on the compression side. Accordingly, large prior art offsets, i.e., offsets in which the crankshaft must rotate at least 20 degrees past a pistons top dead center position before the piston can reach a firing position, have not been utilized, disclosed or taught. As a result, the relatively large offsets required to substantially align peak torque to peak combustion pressure cannot be accomplished with Lee's invention.
Variable Compression Ratio (VCR) engines are a class of prior art CI engines designed to take advantage of varying the compression ratio on an engine to increase efficiency. One such typical example is disclosed in U.S. Pat. No. 4,955,328 to Sobotowski. Sobotowski describes an engine in which compression ratio is varied by altering the phase relation between two pistons operating in cylinders interconnected through a transfer port that lets the gas flow in both directions.
However, altering the phase relation to vary compression ratios impose design requirements on the engine that greatly increase its complexity and decrease its utility. For example, each piston of the pair of pistons must reciprocate through all four strokes of a complete four stroke cycle, and must be driven by a pair of crankshafts which rotate through two full revolutions per four stroke cycle. Additionally, the linkages between the pair of crankshafts become very complex and heavy. Also the engine is limited by design to CI engines due to the higher compression ratios involved.
Various other relatively recent specialized prior art engines have also been designed in an attempt to increase engine efficiency. One such engine is described in U.S. Pat. No. 5,546,897 to Bracken entitled “Internal Combustion Engine with Stroke Specialized Cylinders”. In Brackett, the engine is divided into a working section and a compressor section. The compressor section delivers charged air to the working section, which utilizes a scotch yoke or conjugate drive motion translator design to enhance efficiency. The specialized engine can be described as a horizontally opposed engine in which a pair of opposed pistons reciprocate in opposing directions within one cylinder block.
However, the compressor is designed essentially as a super charger which delivers supercharged gas to the working section. Each piston in the working section must reciprocate through all four strokes of intake, compression, power and exhaust, as each crankshaft involved must complete two full revolutions per four-stroke cycle. Additionally, the design is complex, expensive and limited to very specialized CI engines.
Another specialized prior art design is described in U.S. Pat. No. 5,623,894 to Clarke entitled “Dual Compression and Dual Expansion Engine”. Clarke essentially discloses a specialized two-stroke engine where opposing pistons are disposed in a single cylinder to perform a power stroke and a compression stroke. The single cylinder and the crowns of the opposing pistons define a combustion chamber, which is located in a reciprocating inner housing. Intake and exhaust of the gas into and out of the combustion chamber is performed by specialized conical pistons, and the reciprocating inner housing.
However, the engine is a highly specialized two-stroke system in which the opposing pistons each perform a compression stroke and a power stroke in the same cylinder. Additionally, the design is very complex requiring dual crankshafts, four pistons and a reciprocating inner housing to complete the single revolution two-stroke cycle. Also, the engine is limited to large CI engine applications.
Accordingly, there is a need for an improved four-stroke internal combustion engine, which can enhance efficiency by more closely aligning the torque and force curves generated during a power stroke without increasing compression ratios substantially beyond normally accepted design limits.